Variable displacement hydraulic assembly



July 17, 1962 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 1 awe,

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VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY '7 Sheets-Sheet 2 B mm W A m A U m E mm w w 0 o 0 T A E W on 2 Q Q \L mm B 9 mm 4N A Hi 2 3 n. III L m flw fimw m+ ||||||.|.l|1||| I ||r H '1 I. A m mm 40 A 2 mm i E. m 6 mm @NN 3 m m Q QA w 6 mm, 0% w a Nu o A E .mm & w l 0 vm A, 0% HN w o o 0 o o \L m m July 17, 1962 Filed July 9, 1957 July 17, 1962 L. T. HARRIS 3,044,409

VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Filed July 9, 1957 7 Sheets-Sheet :5

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7 Sheets-Sheet 4 Filed July 9, 1957 INVENTOR. LEE T HARRIS July 17, 1962 L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY 7 Sheets-Sheet 5 Filed July 9, 1957 INVENTOR. LEE T. HARRIS O fizufa an L. T. HARRIS VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY July 17, 1962 Filed July 9, 1957 7 Sheets-Sheet 6 INVENTOR.

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VARIABLE DISPLACEMENT HYDRAULIC ASSEMBLY Filed July 9, 1957 '7 Sheets-Sheet 7 TO ENGINE VACUUM TO BATTERY 68 TO IGNITION SW. 1

T0 BAND BRAKE 1 SOLENOID SERVO-VALVE TO BAND BRAKE SOLENOID SERVO-VALVE NEUTRAL 120 BRUSH BRUSH RELAY R RELA F INVENTOR- BY LEE I HARRIS United States Patent Filed July 9, 1957, Ser. No. 670,814 9 Claims. (Cl. 103120) This invention relates to hydraulic pump mechanisms and more particularly to a variable displacement pump mechanism that may be used in a driving and driven configuration to comprise a unique hydraulic transmission system.

For some years new variable displacement pumps have been used in hydraulic transmission systems for coupling a prime mover to a load such as in automobiles and other industrial applications. These transmissions and pumps have generally been satisfactory for various applications but have had certain disadvantages such as limited inputoutput speed ratios requiring special gear trains, high frictional and hydraulic losses resulting in low efliciency transmissions, costly and expensive machining of highly stressed and critical parts resulting in high cost limited installations. According to the present invention I have discovered a variable displacement hydraulic pump and transmission system that permits infinite variation in input-output speed and torque ratios that is dynamically and statically balanced so as to minimize frictional forces and that still may be manufactured largely by castings in a cheap and simple manner.

Accordingly it is an object of the present invention to provide a variable displacement pump mechanism that is infinitely variable in displacement. It is another object of the present invention to provide a transmission system that is infinitely variable in input-output speed ratio throughout the design limits thereof. It is another object of the present invention to provide a variable pump mechanism that is dynamically and statically balanced so as to minimize frictional forces therein. It is another object of the present invention to provide a transmission system that is extremely flexible in application and efiicient in operation. It is another object of the present invention to provide a transmission system wherein the driving pump may be located remotely from one or more driven pump mechanisms. It is another object of the present invention to provide a variable displacement hydraulic pump mechanism that may be combined with a corresponding mechanism to provide a transmission that rotates together at a one-to-one speed ratio without external gearing or locking. It is another object of the present invention to provide a transmission system that requires no gear train or other speed control mechanism to connect it between a prime power source and a load. It is a still further object of the present invention to provide a variable displacement pump and transmission mechanism that is highly eflicient, extremely simple to operate and economical to construct. These and other and further objects will be in part apparent and in part pointed out as the specification proceeds.

In the drawings:

FIGURE 1 is an axial section of the transmission with the rotatable inner portion shown at a position wherein the pitch axis of the driving and driven pump divider disks are perpendicular to the plane of the drawing and wherein the transmission is at an input-output forward speed ratio of one-to-one;

FIGURE 2 is an axial section on line II-II of FIG- URE 1 of the transmission except that the divider disk of the driving pump is at the zero pitch position;

FIGURE 3 is a side elevation of the bearing member in the same aspect relationship as shown in FIGURE 1;

FIGURE 4 is a top plan partially broken away of the bearing member of FIGURE 1;

ice

FIGURE Sis a vertical section of the bearing member along the line VV of FIGURE 3;

' FIGURE 6 is an axial section'of the ducting core which fits into the bearing member in the same aspect relationship as shown in FIGURE 4 with certain parts shown in full lines for clarity;

FIGURE 7 is a view similar to FIGURE 6 taken on line VIIVII of FIGURE 6;

FIGURE 8 is a partial sectional view on line VIII-VIII of FIGURE 7;

FIGURE 9 is an end elevational view of the disk support assembly as viewed from the left hand side of FIGURE 2 and FIGURE 10;

FIGURE 10 is a side elevational view of the disk support assembly in the same aspect relationship as shown in FIGURE 2;

FIGURE 11 is a top plan view of the disk support assembly of FIGURE 10;

FIGURES 12 and 12A are from left to right an end and side view of either of the two bushings that fit onto the transverse projections of the disk support assembly;

FIGURE 13 is a diagrammatic drawing of the transmission main oil ducting described by 360 counterclockwise rotation as viewed from the left hand end of FIGURE 1 of the ducting core, starting at top center in respect to FIGURE 1;

FIGURE 14 is a side view of the pump chamber end piece in the same aspect relationship as shown in FIG- URE 1;

FIGURE 15 is atop plan view of the chamber end piece of FIGURE 14;

FIGURE 16 is a vertical transverse section of the pump chamber end piece along the line XVI-XVI of FIG- URE 14;

FIGURE 17 is a right half end view of the frontal en plate of the inner housing as viewed from the right hand side of FIGURE 1;

FIGURE 18 is an axial section of the frontal end plate of the inner housing in the same aspect relationship as shown in FIGURE 1;

FIGURE 19 is an axial section of the frontal end plate of the inner housing in the same aspect relationship as shown in FIGURE 2;

FIGURE 20 is an end view of the intermeshed rotor and divider disk of the driving pump as viewed from the right in FIGURES 1 and 2;

FIGURE 21 is a side view of the driving pump rotor and divider disk of FIGURE 20 shown with the latter in section along line XXIXXI of FIGURE 20;

FIGURE 22 is a partial transverse section of the driving pump rotor and divider disk along the line XXII- XXII of FIGURE 21;

FIGURE 23 is an enlarged view of a rotor vane and a partial circumferential section of the intermeshed divider disk showing the relationship of the two to each other at a phase of the rotation cycle where the maximum angle of incidence occurs in respect to the normal; 7

FIGURE 24 is a partial radial sectionof the dividerdisk along the line O--XXIV of FIGURE 22 With the rotor removed;

FIGURE 25 is a partial radial section of the divider disk along the line OXXV of FIGURE 22 with the rotor removed; 7

FIGURE 26 is an enlarged exploded perspective wew of a divider disk insert and the corresponding urging p FIGURE 27 is a schematic representation of a method for forcibly maintaining a relatively uniform speed relationship of the divider disk in respect to the rotor;

FIGURES 28, 29, and 30 are enlarged top, side, and end views respectively of a rotor vane showing a further alternative method of forcibly maintaining a relatively uniform speed relationship of the divider disk in respect to the rotor; and

FIGURES 31 and 32 are schematic drawings of a possible control system for automatically controlling the input-output speed ratio of the transmission for automotive application.

Referring now to FIGURES 1 and 2, the transmission includes an input shaft 1; an output shaft 2; an outer housing comprising end plate 3 and upper and lower casings 4 and 5 respectively; an inner housing comprising end plate 6, front and rear barrel sections 8 and 9, central section 10, and rear section 12; ducting core 13; a first pump chamber bounded by pump chamber end piece 7, bearing member 11, and disk support assembly 16-16; a second pump chamber bounded by member 1% and rear section 12; a driving pump including a rotor 14 and divider disk 15-15; a driven pump or motor including rotor 17, and divider disk 18-18; over-running clutch U; and a band brake including band 21 and drum The driving pump is infinitely variable in displacement from design maximum forward output through zero output to design maximum reverse output as will be subsequently described. The rotor 14 is mechanically'cowpled to input shaft 1 and has therein six radial vanes 22 (see FIGURE 20) circumferentially spaced 60 apart at the centers with fiat sides and peripheral surfaces which are as surfaces of a common sphere Whose center is the geometric center point of the pump chamber, and a core section whose outer surface 23 is spherically contoured with the sphere center coinciding with the same aforementioned geometric center point of the pump chamber, said surface 23 as the innermost surface or boundary of the pump chamber. Front face surface 24 of chamber end piece 7, rear face surface 25 of bearing member 11 and peripheral surface 26 of assembly 16-16 define the other principal boundaries of the first pump chamber and are shaped so as to provide an oil sealing contact to the 7 end and peripheral surfaces of revolution of the vanes 22 of rotor 14. Thus the six vanes 22 circumferentially divide the pump chamber into six separate and equal volumes of fixed displacement.

Divider disk 15-15 comprises two main sections (see FIGURES 20, 21, and 22); six slots 27 for accommodation of the vanes 22 of rotor 14, recesses 23 for accommodating inserts 29, valve chambers 39 for accommodation of shuttle balls 31, oil channels 32, recess grooves 33, oil channels'34, and means for securely fastening the two main parts 15 and 15 to each other. Said divider disk 15-15 intermeshes with rotor 14 as shown in FIG- URES 20, 21 and 22 and serves to longitudinally divide each fixed volume'between adjacent vanes 22 into two par-ts. As visualized in FIGURE 21, divider disk 15-15 would be assembled on rotor 14 by sliding the part 15 onto said rotor 14 from the left and sliding the other the phase diiference of slots 27 in respect to correspondpart 15' on from the'right, one part, either 15 or 15 containing the inserts 29 with corresponding springs and shuttle balls 31, and the two parts 15-15 being joined by screws 36 as shown in FIGURES 20 and 24 or. in some other suitable manner. Divider disk 15-15 is supported in axial position by contact or proximity of the outer face bearing surfaces 37 of said divider disk' 15-15 with the recessed mating surfaces 37 of disk sup-' port assembly 16-16 (FIGURES 2 and 10) and in 15-15 and rotor 14 is provided by contact or proximity of the spherically contoured inner contact surfaces 23' and outer slot contact surfaces 27 of said divider disk 15-15 with the mating spherically contoured contact surfaces 23 and 22' of said rotor 14 in any supported position of said divider disk 15-15 within the limits of movement of disk supportassembly 16-16 about its transverse axis. Oil sealing contact between the slots 27 and the radial surfaces of vanes 22 is maintained by inserts 29 which are slidably mounted in recesses 28, said inserts 29 being continuously but yieldably urged against vanes 22 by springs 35 or hydraulic pressure channeled into recesses 28 by shuttle balls 31 or a combination of the two forcing means. Shuttle balls 31 actually serve two purposes: (1) by being exposed to the pressure appeering on both sides of divider disk 15-15', between any two adjacent vanes 22, through orifices 39 said shuttle balls 31 are automatically seated against the orifices 39 which are facing the lowest pressure, thereby leaving the orifices 39 which are facing the highest pressure open, and permitting the highest available pressure to be transmitted into valve chambers 30 and then into recesses 28 via oil channels 32 wherein said high pressure acts against the backsides of inserts 29 thereby counteracting the highest pressure appearing on the exposed face surfaces ll of said inserts 2? (FIGURES 23 and 26) (2) by leaving one orifice 3% open at all times, the oil trapped in recesses 28 by inserts 29 is provided an escape when said inserts 29 are forced back into said recesses 28 from extended positions.

In operation it will be understood that when the input shaft 1 is rotated, as for example by a prime power source, that rotor 14 likewise rotates by virtue of its direct coupling to input shaft 1 and that since divider disk 15-15 is mechanically intermeshed with rotor 14, it will likewise rotate with said rotor 14 but will be permitted some circumferential movement relative to rotor 14, said relative movement being restricted to the limits corresponding to the excess clearance between the face contact surfaces 40 of the opposing inserts 29, on each side of a particular slot 27 when said inserts 29 are in their fully retracted positions, as compared to the thickness of vanes 22. Said excess clearance should be at least as great as the maximum phase deviation of corresponding points of rotor 14 and divider disk 15-15 when calulated from the equation tan B=tan A cos P, in which P is the pitch angle of divider disk 15-15, A is any angle of the complete 360 cycle of rotation of rotor 14 with either side of the pitch axis of divider disk 15-15 as the zero degree reference and B is the angle of correspondence of the angle A on said divider disk 15-15 when translated to the plane of rotation of rotor 14-. Therefore, angle B subtracted from the angle A represents ing vanes 22 for various angles of rotation and for various pitch angles at which divider disk 15-15 may be set between zero and maximum design limits. In order to arrive at the minimum design clearance for vanes 22 l in slots 27, angle A should be taken at angles of 45,

radial position by rotor 14. Disk support assembly 1eject to control in its movements about the axis as will be describedherein. v

In operation oil sealing action between divider disk 135, 225 and 315, and angle P at the maximum design pitch angle of divider disk 15-15. Maximum angle P is 15 in the illustrated embodiment which causes the maximumphase difference between rotor 14 and divider disk 15-15 to beapproximately :1; therefore the total clearance between face contact surfaces of opposing inserts 29 when in their fully retracted positions in grooves 28 should be at least equal to'the thickness of vanes '22 plus r sin 2, where r is'the radius from the geometric center of the pump housing.

Although inserts 29 will perform most of the yielding to the phase variation between rotor 14 and divider disk 15-15, the frictional resistance to rotation against dividerdisk 15-15 will tend to cause said divider disk 15-15 to lag behind rotor 14 to the limits imposed by the excess clearance between opposed inserts 29 which 5 will tend to cause cyclic variations in the speed of said divider disk 15-15 in respect to the speedrof rotation of rotor 14. FIGURE 27 illustrates a means whereby a more nearly constant speed rotation of divider disk 15-15 in respect to rotor 14 may be assured, when the transmissionis operating at an input-output speed ratio of one to one. A groove 41 is provided in the peripheral surface 22 of each vane 22, the sides of which are shaped to provide contact to at least one point of a mating projection 42, located at the peripheral center of each slot 27 in divider disk 15-15, at all phase angles of rotation between 315 to 45 and 135 to 225, with the zero degree reference considered to be in coincidence with one side or the other of the transverse pitch axis of divider disk 15-15.

A further alternative means whereby more nearly constant speed rotation of divider disk 15-15 in respect to rotor 14 may be assured (FIGURES 28, 29, and 30) is to shape the sides of vanes 22 so as to keep one or both of the opposing inserts 29 in any one slot 27 fully retracted into their respective recesses 28 at all phases of cyclic rotation between the angles of 315 to 45 and 135 to 225 with the zero degree reference again chosen to coincide with either side of the transverse pitch axis of divider disk 15-15.

By inspection of the drawings, it will be apparent that divider disk 15-15' operating in a complementary manner with rotor 14 essentially divides the pump chamber into two pumping sections (one on each side of divider disk 15-15) which operate in a diametrically opposite fashion to each other with each section having its own input-output ports, represented by the numbers 43 and 43 for the front pump section and numbers 44 and 44' for the rear pump section as shown in FIGURE 4 and as shown schematically in FIGURE 13. The ports 43, 43, 44 and 44' should each have a circumferential length of approximately 120, leaving an un-cut-away circumferential section of approximately 60 in length between each port which will serve to trap the fluid between adjacent vanes 22 at maximum or minimum points as the case may be and thereby to prevent unrestricted circula- 'tion of oil between input and output ports which would tend to bypass the pump. If the rotor employs more or less than six vanes as shown in the drawings, the ports should of course be of appropriate circumferential length corresponding to the number of vanes.

Referring now to FIGURE 1 it is obvious that when the plane of rotation of divider disk 15-15' is completely vertical (zero pitch angle) represented by alignment with the letters N-N, the volume of fluid entrapped between adjacent vanes 22 and the surfaces of the pump chamber remains constant when rotor 14 and divider disk 15-15 rotate and that therefore all of the fluid contained within the chamber rotates with the rotor with no fluid input or output through ports 43, 43, 44 and 44. It should be equally obvious that if the pitch angle of divider disk 15-15 is at any value other than zero that the individual fluid volumes between adjacent vanes 22 on each side of divider disk 15-15 will vary in a sine manner as rotor 14 progresses from through 360 for each rotation cycle, the total change between minima and maxima volumes being a direct function of the pitch angle of divider disk 15-15. Therefore a low compressibility fluid such as oil when contained therein will be forcibly circulated into and out of the pump chamher through ports 43, 43, 44 and 44 in amounts proportional to the pitch angle of divider disk 15-15. In the illustrated embodiment, maximum forward circulation is achieved when the plane of rotation of divider Examination of the underlying principles of operation will disclose that insofar as rotor 14 is concerned, no appreciable unbalanced forces in respect to the housing occur but that the hydraulic forces on divider disk 15- 15 would be severely unbalanced if not compensated for. It is for said purpose of compensation that recess grooves 33 are provided around each peripheral side of divider disk 15-15 (FIGURES 20 and 21) with oil channels 34 providing hydraulic communication between said recess grooves 33 and the opposite side of divider disk 15-15 in the pump chamber. The effective area of recess groove 33 is chosen so that said area multiplied by the radius of the effective center of fluid pressure of said recess groove 33 is equal to the effective area of the part of divider disk 15-15 contained inside the pump chamber adjacent vanes 22 multiplied by the radius of the effective center of fluid pressure of said second area. In this way any hydraulic forces appearing on the inside areas of divider disk 15-15 will be dynamically counteracted by proportionate forces created by transmis: sion of pressure through oil channels 34 to recess grooves 33 located on the opposite sides. The obvious purpose of providing hydrostatic balance for divider disk 15-15' is to reduce the frictional resistance to rotation of said divider disk 15-15 in disk support assembly 16-16 occurring as a result of hydraulic loading, and thereby to enhance the efliciency of power transfer and decrease mechanical wear. Another method for reducing the friction, in lieu of hydrostatic balance as described above, is to support divider disk 15-15 on ball or tapered roller bearings within disk support assembly 16-16. A third method might be a combination of the above two methods.

The driven pump or motor, hereinafter to be referred to as the driven pump, in the illustrated embodiment is essentially identical to the driving pump except that divider disk 18-18 has a fixed plane (pitch angle of 15) of rotation about its axis by virtue of its being supported in the inner housing and accordingly is a fixed displacement pump or motor. It will be apparent, however, that the driven pump may be of the variable displacement type also =and that said driven pump whether of the fixed or variable displacement type may be similar to -or different from the driving pump in basic design. The rotor 17 with six vanes 22A, divider disk 13-18 and associated pump chamber acting in cooperation with each other form the essential elements or" the driven pump, the operation of which will be fully understood by reference to the above description of the driving pump.

connects ports 43 and 44 of the driving pump to ports 46 and 47 of the driven pump.

It will be apparent that when input shaft 1 is rotated clockwise when viewing FIGURES '1 and 2 from the left, and that when divider disk 15-15 is set at a pitch angle greater than Zero toward alignment with the letters F-F of FIGURE 1, fluid will be expelled from the chamber of the driving pump through ports 43 and 44 into channel 45 from which said fluid will be forced into the chamber of the driven pump through ports 46 and 47 thereby causing rotor 17 to rotate in the same clockwise sense as rotor 14, causing said rotor 17 at the same time to expel a like quantity of fluid from the chamber of the driven pump through ports 46 and 47 into channel 45 from which said like quantity of fluid will be returned to the chamber of the driving pump through ports 43 and 44. Fluid flow under the above conditions is shown schematically by the solid arrows in FIGURE 13. When divider disk 15-15 is set at a pitch angle greater than zero toward alignment with the letters R-R of FIG- URE 1 and rotation of inputshaft 1 is in the same clockwise sense, the flow of fluid isrreversed and would be as 7 shown schematically by the dashed arrows in FIGURE 13, in which case output shaft 2 would rotate counterclockwise or in the opposite direction to the direction of rotation of input shaft 1.

It is also apparent that regardless of the direction of rotation of input shaft 1, with divider disk 15-15 pitched toward alignment with the letters FF in FIG- URE l, the output shaft 2 will rotate in the same direction as input shaft 1, and with divider disk 15-15 pitched toward alignment with the letters RR in FIG- U-RE 1, rotation of output shaft 2 will be opposite to that of'input shaft 1.

It will be understood that the ratio of speed of the input shaft 1 to the speed of the output shaft 2 will be inversely proportional to the ratio of the displacement of the driving pump to the displacement of the driven pump and that the output-input torque ratio will be directly proportional to the input-output speed ratio, neg ecting frictional losses. In the illustrated embodiment of the transmission, a one to one forward ratio of speed between input shaft 1 and output shaft 2 occurs when the plane of rotation of divider disk 15-15 is in alignment with the letters FF in FIGURE 1.

Pitch control of divider disk 1515 is effected by control of the movement of disk support assembly 16-l6' about its transverse axis by permitting oil under pressure to enter either control chamber A through channel 43 or control chamber B through channel 49 and to exit from the other (see FIGURE 1). Disk support assembiy 1616' has projections or vanes Si) (FIGURES 9, I0, 11) which together with cylindrical surfaces 51 (FIG- URES 1'7, 18, 19) in end plate 6 and the spherical peripheral surfaces 51"of control chambers A and B (FIG- URE 1) forms a fluid seal to prevent the unrestricted flow of oil between control chambers A md B. The axial center line of said cylindrical surfaces 51 coincides with the transverse axis of rotation of disk support assembly 16-16. Vanes 52 are similar to vanes 50 but are not necessarily designed to provide oil sealing contact with any surfaces of the housing, the essenn'al purpose of said vanes 52 being to balance the weight of vanes in respect to the longitudinal axis of the transmission when the inner housing is rota-ting, although said vanes 52 also serve to strengthen the disk support assembly 1l616 at points where needed. 7

To increase the input-output forward speed ratio or to decrease the input-output reverse speed ratio, oil under pressure is permitted to enter control chamber B through channel 49 from the control pressure source and a like amount of fluid is permitted to exit from control chamber A through channel 48 and to return to the sump or inlet of the control pressure source, which action causes disk support assembly 1616 to move in the direction toward alignment of the transverse center plane of said divider disk 1616' with the letters .FF of FIGURE 1. To decrease the input output forward speed ratio or to in crease the input-output reverse speed ratio, oil under pressure is permitted to enter control chamber A through channel 48 and to exit from control chamber B through channel 49 and to return to the sump or inlet of the control pressure source, which action causes disk support assembly to move in the direction toward alignment of the transverse center plane of said divider disk i1616 with the letters RR of FIGURE 1. Annular grooves 48' and 49' are provided in end piece 3 of the outer hous: ing as a means of effecting hydraulic communication between a control system located externally to the inner housing and channels 48 and 49 respectively under conditions when the inner housing is'rotating as well as when said inner housing is held motionless in respect to the outer housing as shown in FIGURES 1 and 2.

In order to obtain a mechanical indication of the transmission input-output speed ratio which might be required for a follow-up type of control system, a ratio indicating linkage is provided as shown in FIGURE 6. This linkage starts with a pair of gears 53 and 53 fixed on shaft 55 and mounted so that gear 54- meshes with a circular gear segment 56 (FIGURE 11) on disk support assembly 16-16 and gear 54 meshes with gear rack 57 mounted in sleeve 58 inside ducting core 13. When disk support assembly -1616' rotates about its transverse axis a proportionate angular rotation, corresponding to the gearing ratio between said gear 53 and said circular gear segment 56, is induced in gear 54 causing gear rack 57 to move longitudinally together with sleeve 58 to which it is mechanically fixed. Shaft 59 is threaded into sleeve 58 at one end (and'therefore moves longitudinally with said sleeve 58) and has fixed in position on the other end of shaft 59 by retaining nut 61 a bearing assembly 60 which is slidably mounted in output shaft 2 (FIGURE 2). An indicating sleeve 62 having thereon a. rim 66 is slidably mounted on output shaft 2 and mechanically fastened to bearing assembly 60 by screws 63 which are free to move longitudinally in slots 64, said indicating sleeve 62 thereby being caused to move longitudinally as a unit with bearing assembly 60. A lever assembly consisting of two arms 65 and 65', the extremities of which bear against rim 66 of indicating sleeve 62 fixed on a shaft 67 to which a third arm 68 is rigidly fastened extends the linkage away from the output shaft to the control mechanism. Shaft 67 is trunnioned on a transverse axis so that its center line is the pivot point for the lever assembly which is caused to pivot about its axis to follow the movement of indicating sleeve 62 by spring 85 or other equivalent means. A rod 69 is pivotally mounted on the end of arm 68 so as to translate the rotative movement of said arm 68 into longitudinal motion for utilization by a transmission control means. Rotative motion could just as well be imparted to the transmission control means by appropriate gearing arrangement from shaft 67 of the lever assembly if said rotative motion were more suitable for use by the control means.

When disk support assembly 1616' rotates about its transverse axis a proportionate angular rotation, corresponding to the gearing ratio between said gear 53 and said circular gear segment 56, is induced in gear 54 cansing gear rack 57 to move longitudinally together with sleeve 58 to which it is mechanically fixed. This movement is in turn transmitted through the linkages described until rod 69 moves the indicating mechanism to the corresponding indication.

Bypass control piston 70 (FIGURES 6 and 7), forming the end of shaft 59 which attaches to sleeve 58, contains an annular groove 71, the purpose of which is'to bypass a nominal amount of oil from the high pressure channel in ducting core 13 via orifices 72 and 72' to the low pressure channel in said ducting core 13 when a load connected to output shaft 2 is being accelerated from the motionless state in either the forward or reverse direction. Bypass control piston 7% acts in cooperation with orifices 72 and 72 to form'a progressively smaller restriction to the bypass oil path as divider disk 15-45 is increased in pitch in either the forward or reverse direction from the zero pitch angle up to a predetermined pitch angle at which point the bypass path will be completely obstructed by the smooth cylindrical surface of the bypass control piston. By this means, torque in varying degrees may be applied to a motionless load without stalling the power source, thereby permitting smoother starts. For some forms of application of the transmission a bypass valve obviously would not be required. v

Shuttle valve 73 (FIGURES 6 and 7) may be mounted in the ducting core 13 by threaded means as shown and comprises a valve housing 74 and shuttle spool 75. The valve on one end is exposed to fluid channel 45 which is normally the high pressure channel and on the other end is exposed to fluid channel 45' which is normally the low pressure or return channel. The normally highpressure and return channels 45 and 45' may alternate with each other pressure-wise, depending upon the mode of operation of the transmission, however, the shuttle spool 75 is always forced against the valve seat on the high pressure side thus closing off the high pressure from the center of ducting core 13; on the other hand the low pressure side is always in hydraulic communication with the center of the ducting core 13 so that except for the pressure drop due to oil flow through the communication channeling, the low pressure or return prime channel 45 or 45' will be at the same fluid pressure as the center of ducting core 13. By this device a charging pressure may be applied to the return fluid channel 45 or 45' through channel 76, located at the center of input shaft 1 (FIGURES l and 2), to discourage any tendency toward cavitation by pumping action of the driving and driven pumps and also to make up any leakage of oil from the inner housing of the transmission. Annular groove 77 is provided in end piece 3 of the outer housing as a means of effecting hydraulic communication between an oil pressure source located externally to the inner housing and channel 76 under conditions when the inner housing is rotating as well as when said inner housing is held motionless in respect to the outer housing as shown in FIGURES 1 and 2.

Referring to FIGURE 2 it will be apparent that spring loaded piston 78 may be a pressure regulating device which will maintain a certain charging pressure depending upon the spring characteristics and certain other factors when oil is forcibly circulated through channel 79 at the center of shaft 59. Oil flow would be through oil chan- 7 6 from annular groove 77 into the center of ducting core 13 from where it would be transmitted through channel 79 located at the center of shaft 59 to the face of piston 78, which would cause the piston to move toward the right in FIGURE 2 until orifices 89 were exposed to the extent necessary to pass the volume of oil being circulated. Oil would flow through orifices 86 into the hollow space in output shaft 2 partially occupied by bearing assembly 69 from Where said oil would then pass through slots 64 and eventually along the lower side of the outer housing where it would be returned to the sump through channel Si in end piece 3 (FIGURE 1). Circulation of oil in this manner may also provide a means of transferring heat from the transmission proper to the sump or to a heat exchanger; however, under most conditions the power loss in the transmission will be sufiiciently low as to render the use of a large capacity heat exchanger unnecessary.

The inner housing comprising pa-rts aforementioned is journaled at the input shaft end on bearings 82 and at the output shaft end on bearings 83, said inner housing being disposed to rotate under certain conditions as described herewith. When operating at an input-output forward speed ratio of one to one, the reaction forces imparted to said inner housing through divider disk 15-15 are equal and opposite to the reaction forces imparted to said inner housing through divider disk -18-'i8, said forces effectively cancelling each other, thus permitting internal frictional forces to rotate said inner housing without the necessity for fluid locking. Assuming some leakage of fluid which bypasses the normal circulatory paths provided in the transmission, output shaft 2 will experience a slippage in speed relative to the speed of input shaft 1, in which case it is obvious that the inner housing cannot rotate at synchronous speed with both input shaft 1 andou-tput shaft 2. Dynamic forces acting on divider disk 15- 15 of the driving pump and divider disk 1818' of the driven pump when the two are in asynchronous motion cause the inner housing to rotate at a speed intermediate between the speeds of the input and output shafts due to-the uniformity of pressure in directly communicative spaces.

Circulation of oil with the inner housing in synchronous rotation will be near zero; however, this will not impair torque transfer inasmuch as the torque transfer is a function of the net resultant hydraulic forces acting on vanes 1Q 22 and 22A of the rotors 14 and 17 and not a function of the oil transfer.

When the input-output forward speed ratio is greater than one-to-one, and when the power source is supplying torque to input shaft 1, greater torque is applied to driven pump rotor 17 than is applied to driving pump rotor 14, resulting in a net reverse reaction force acting on the inner housing through the respective divider disks 18-18 and 16-46 which tends to cause said inner housing to rotate in the opposite direction to the rotation of the input and output shafts. The function of over-running clutch 19, shown diagrammatically in FIGURES 1 and 2, is to prevent the reverse rotation of the inner housing but to permit free forward rotation.

When the input-output reverse speed ratio is greater than oneato-one, and when the power source is supplying torque .to input shaft 1, greater torque is applied to driven pump rotor 17 than is applied to driving pump rot-or 14, and the net reaction force acting on the inner housing in this case tends to cause said inner housing to rotate in the forward direction which is undesirable inasmuch as said forward rotation would cause an effective reduction in the output-input torque-ratio. Band 21, when caused to contract around the drum 2%} with adequate force by hydraulic servo means or other suitable means, prevents forward rotation of the inner housing.

Under conditions when the impelling force is transminted v1a output shaft 2 to rotor 17, thence to rotor 14 and then via input shaft 1 to the power source which in this case becomes the load, as for example when a vehicle is descending a grade, and it is desired to increase the load by forcing .the power source to a higher speed by increasing the transmission input-output speed ratio, hydraulic forces tend to rotate the inner housing in the forward direction, which rotation if permitted to occur would lower the input-output speed ratio; therefore it is again desirable for the inner housing to be held against forward rotation by constriction of the band 21 around the drum 2th Shield 34 is provided for the purpose of reducing windage losses when the inner housing is rotating and is de signed to cover the circumferentially asymmetrical portion of the inner housing.

Possibilities for many different configurations of control systems exist for adapting a transmission of this type to automotive, industrial or other applications. FIGURE 31 is a simplified schematic drawing of a control system which might be used to automatically control the inputoutput speed ratio of the transmission in automotive vehicles as a function of the output shaft speed and torque demand, or input shaft speed under certain conditions. The control system herein described is jointly electrical, mechanical and hydraulic in operation.

The hydraulic portion includes an oil pressure source 109 and a follow up valve assembly'ltll containing a valve spool 102 which is axially slid-able in valve sleeve Hi3, said valve sleeve 103 being slid-ably mounted in cylinder 104. The lands 105 and 196 on the valve spool 102 are of such width and spacing so as to completely occlude oil inlet channel 107 and oil return channel 198 when aligned as shown in FIGURE 31, but to permit circulation of oil from the pressure source 104 through the control passages and back to the sump when said valve spool 102 is displaced axially in either direction in respect to valve sleeve 183. When valve spool 102 is moved to the left in FIG- URE 31, oil under pressure from inlet channel 107 is permitted to enter control chamber B through oil channel 189 and oil is permitted to return from control chamber A to return channel 168 through oil passage as indicated by the solid arrows. By this means disk support assembly 1616' moves in 'the direction of increasing forward pitch or decreasing reverse pitch caus ing arm 68 of the ratio indicating linkage to move in the direction of the solid arrow. This action continues until inlet channel 107 and return channel 108 are realigned tation rotor.

1 i with lands 1G5 and 106 at which time oil circulation into control chamber B and out of control chamber A is blocked. If valve spool 162 is moved to the right in FXGURE 31 oil is permitted to enter control chamber A under pressure and to exit from control chamber B as indicated by the dotted arrows. This causes disk support assembly 16-16' to move in the direction of decreasing forward pitch or increasing reverse pitch causing arm 53 in this case to move in the direction of the dotted arrow. This action continues until inlet channel 107 and return channel 108 are again realigned with iands 1195 and 1% at which time oil circulation into control chamber A and out of control chamber B is blocked. It is therefore ap parent that when valve spool 162 is moved axially in either direction, that disk support assembly 1616 is caused by hydraulic means to follow said axial movement of valve spool 102 by a corresponding angular amount in the direction which will tend to maintain alignment of lands 105 and 1% with inlet channel 107 and return channel 108.

An electric motor 111 is provided to effect axial movement of valve spool 102 through cooperation of the threaded parts of shaft 112 and spool shaft 113. Spool shaft 113 should be designed for axial movement only. Motor 111 is controlled electrically by commutator-brush assembly 114 containing four pairs of commutatorbrush rotors as shown schematically in FIGURE 32; viewing from left to right, one pair for Drive, one pair for Neutral, one pair for Low and one pair for Reverse. Essentially the purpose of commutator-brush assembly 114 is to compare the actual input-output speed ratio of the transmission at any instant with what the ratio should be as a function of transmission output shaft speed, torque demand, and selected speed range and to control motor 111 in such a way as to cause said motor 111 to drive vflve spool 1112 to the position which will correct any discrepancy.

Shaft 115 is geared at one end to spool shaft 113 and is directly coupled to the Drive, Neutral, and Low brush arms or rotors and is coupled through reverse gearing (not shown) to the Reverse brush arm and thereby causes angular movement of said brush arms when spool shaft 113 moses axially. Shaft 116 is geared at one end to control rod 117 and is coupled by suitable means to the Drive, Low, and Reverse commutation rotors so as to provide angular rotation in direct proportion to axial movement of said control rod 117 up to certain predetermined limits for each commu- For example, the limit of travel for the Drive commutation rotor might be adjusted to correspond to an input-output speed ratio of one-to one for the transmission, whereas the limit of travel for the Low and Rcver-se'commutation rotors might arbitrarily be 'adjusted'to correspond to an inputcutput speed ratio of two-to-one for the transmission. The Neutralcommutation rotor is fixed in position corresponding to an input-output speed ratio of infinity, or zero output for p the transmission driving pump.

Selector switch 118 may be designed to permit manual selection of any one of the four switch contacts of the provide electrical contact with the brush arm of the one selected. Only the selected pair of commutator-brush rotors may effect control of motor 111. In operation, when control actioncauses the ring F of the selected pair of commutator-brush rotors to come in contact with the brush, a closed circuit is established betweenthe posi ive. and negative poles of the vehicle electrical system via the selector switch 118, brush. arm, ring F and the Winding of relay F. Relay F then establishes electrical connection to motor 111to cause said motor 111 to rot-ate in.

the direction which willcause valve spool 1&2 to move corresponding pairs of commutator-brush rotors and to 1?; gap position 121) (FIGURE 32) between the rings'F and R thus opening the electrical circuit which action causes the motor to stop. When control action causes the ring R of the selected pair of commutator-brush rotors to come in contact with'the brush, a closed circuit is established between the positive and negative poles of the vehicle electrical system via the selector switch 118, brush arm, ring R and the winding of relay R. Relay R then establishes electrical connection to motor 111 to cause said motor 111 to rotate in the direction which will cause valve spool 192 to move to the right viewing FIGURE 31 until the gear 119 and shaft .115 are rotated clock-wise by an amount which will cause the brush arm to move the brush to the gap position 120 between the rings F 'and R thus opening the electrical circuit which action causes the motor to stop.

Since the Drive, Neutral, and Low brush arms are directly coupled to shaft 115, rotation of said brush arms are naturally'in the same direction as for shaft 115; however, inasmuch as the Reverse brush arm is coupled to shaft 115 through reverse gearing, rotation of said Reverse brush arm is in opposition to the rota tion of shaft 115. The reverse gearing of the Reverse brush arm together with the oppositely oriented F and R rings permits the Reverse commutation rotor to move counter-clockwise in respect to the Neutral reference position the same as for the Drive and Low commutation rotors and to provide the same transmission input-output ratio control characteristics in the Reverse speed as in the forward Low speed. The Neutral reference position for the commutation rotors is defined by vertical orientation of the gap position 129 as shown by the commutation rotor for Neutral in FIGURE 32. FIGURE 32 shows the selector switch 118 in the Low position in which a forward input-output speed ratio for the transmission is indicated by a counterclockwise displacement of the Drive," Neutral, and Low brush arms in respect to thevertical' Neutral reference. With the commutation rotors in the same position, it is apparent that if selector switch 118 was put in Reverse position that the servo action previously described would cause valve spool 182 to move to the position representing a reverse transmission speed ratio equal to the forward speed ratio as indicated in FIGURE 32 and the Reverse brush arm would then be at the same counter-clockwise displacement angle in respect to the Neutral reference as for the Drive, Neutral and Low brush a ms shown in FIGURE 32 and that the latter brush arms would then be at the same relative position as shown for the Reverse brush in FIGURE 32.

From the foregoing description it will be understood that input-output speed ratio control of the transmission in either forward or reverse output speeds is effected by control rod 117. As a means of causing said control rod 117 to move as a function of output shaft speed and torque demand or input shaft speed, an output speed gov ernor 121, a vacuum modulator 122 and an input speed governor 123 are provided.

7 Governor 121 may be a centrifugally operated device geared to or otherwise coupled to the transmission output shaft 2 and coupled to control rod 117 as shown in FIG- URE 31 to cause said control rod 117 to move axially to the left, viewing FIGURE 31, as a function of output shaft speed. Vacuum modulator 122 may be coupled to governor 121 in a suitable manner to oppose the action of said governor 121 as a function of engine manifold vacuum and thus provide higher input-output transmission ratios with decreasing manifold vacuum or increasingly open carburetor throttle positions. With output shaft 2 motionless such as would occur when a vehicle to the left viewing FIGURE 31; until the gear 119 and is at a standstill, governor shaft 124 may be set at such a positiion as to cause the transmission pump to be at the zero output or neutral position when spring 125 is holding control rod 117 toward the right, in FIGURE 31, to

. 13 the limit imposed by shoulder 126 resting against the lip of sleeve 127.

Governor 122 may be a c-entrifugally operated device geared or otherwise coupled to the transmission input shaft 1 and coupled to control rod 117 as shown in FIG- URE 31 to cause said control rod 117 to move axially to the left viewin'g FIGURE 31 as a function of input shaft speed. Governor 122 and associated coupling linkage may be designed such that at a nominal input shaft speed, as for example the idling speed of a gasoline engine, shaft 128 will remain at the extreme right hand position, in FIGURE 31, but that at increasingly higher speeds of input shaft 1, shaft 128 will move to increasingly more left hand positions, viewing FIGURE 31, causing control rod 117 to move leftwardly with said shaft 128 against the tension of spring 125 until a predetermined maximum left hand position is reached, for instance corresponding to a transmission input-output ratio of four-to-one, at which point shaft 128 will be held against further leftwardly movement; however, control rod 117 will not be restrained to further leftwardly movement resulting from leftwardly movement 'of sleeve 127. This arrangement will permit smooth starts for vehicles at a standstill, when selector switch 118 is in the Drive, Low, or Reverse positions by permitting governor 122 to exercise control up to a predetermined limit, for instance corresponding to a vehicle road speed of five miles per hour, and permitting governor 121 in conjunction with a torque-demand sensing element such as vacuum modulator 123 to effect transmission control at higher road speeds. As described heretofore, bypass control piston 70 will prevent positive circulation of oil between the driving pump and driven pump of the transmission when the input-output speed ratio is above a predetermined value in forward or reverse output speeds such as would be the case when an automotive vehicle is being accelerated from a standstill. I

v A control means (not shown) of any conventional form may be provided for constricting band brake 20 around drum 21 when either the Low or Reverse positions are selected on the selector switch 118. An interlock switch may be provided to prevent starting of the engine except when disk support assembly 16-16 is at the zero pitch or zero output position.

The control system herein described is suited for automatic operation of the transmission for automotive application without any additional fluid coupling interposed in the power train. The design of the transmission and control system for use with a slippage type of fluid coupling interposed between input shaft 1 and the engine crankshaft could difter in several details.

An 'obviousconstructional variation in the illustrated embodiment would be the physical separation of the drivin-g pump and driven pump with each having its own housing and hydraulically connected by lengths of tubing or conduit. Such a variation would be equivalent to cutting the transmission at line X-X of FIGURE 1 into two parts, each part then being adapted to make appropriate to make appropriate connections with the interconnecting high pressure and return tubes or conduits and removing the outer housing. For instance, by locating the driving pump near the engine and the driven pump near the rear axle of an automobile, the need for the conventional drive shaft and associated universal joints would be eliminated thus making possible a flat floor design for automobiles. By providing two driven pumps or one for each rear wheel, the conventional differential gearing couldbe also eliminated.

A further possibility would be to integrally combine the driven pumps with the wheels.

Constructional variations such as four wheel drive and front wheel drive in lieu of rear wheel drive could more readily be implemented with a hydraulic system than with conventional mechanical power coupling means. The possibility also exists for combining the braking means with the propelling means.

While there is given above a certain specific example of this invention and its application in practical use, it should be understood that this is not intended to be exhaustive or to be limiting of the invention. On the contrary, this illustration and explanation herein are given in order to acquaint others skilled in the art with this invention and the principles thereof and a suitable mannet of its application in practical use, so that others skilled in the art may be enabled to modify the invention and to adapt and apply it in numerous forms each as may be best suited to the requirement of a particular use.

I claim:

1. A variable displacement pump comprising in combination an outer housing having formed therein by nonrotating end plates an inner pumping chamber and an outer control chamber; a work input shaft rotatably mounted in said housing and extending into said pumping chamber; a rotor member operatively connected to said shaft and positioned within said pumping chamber; a plurality of vanes mounted on said rotor; a divider disk member having a corresponding plurality of slots therein; a non-rotating disk support assembly mounted in said outer housing for pivotal motion about an axis disposed at right angles to the axis of said rotor; said disk support assembly having thereon a pair of flanges positioned to form the dividing wall between said inner pumping chamber and said outer control chamber and an annular groove therein disposed outwardly of said pumping chamber adapted to receive said divider disk; said divider disk and disk support serving to divide said inner and outer chambers into two areas of complementary variable capacity compartments; said disk being operatively driven from said rotor by co-operative interaction between disc slots and rotor vanes; a plurality of slots formed in said rotor; fluid ducting means interconnecting said slots for controlling the fiow of fluid in said inner compartments; secondary fluid ducting means interconnecting said outer compartments to control the pivotal position of said disk and support. I

2. A variable displacement pump comprising an outer housing member, a plurality of inner chamber end mem' bers mounted within said housing to form the planar end faces of an inner truncated sphere chamber therein, a first rotor member mounted within said inner chamber and having a plurality of vanes disposed thereon, a dividerdisk assembly pivotally mounted in said outer housing member about an axis thereof, said axis being located at right angles to the axis of said first rotor adjacent the midpoint thereof, a circular divider disk rotatably mounted about the periphery thereof in said disk assembly, a pair of spherical chamber-forming members fixed to said divider-disk assembly adapted to cooperate with said inner chamber end members to form the curved surface of said inner chamber, disposed about the periphery of said divider disc, said divider-disk assembly dividing said outer housing into two substantially equal spaces, a pair of secondary vanes dividing one of said spaces into upper and lower control chambers, oil inlet and outlet means for introducing and withdrawing oil from said control chambers, said divider-disk having therein a plurality of slots adapted to intermesh with said rotor vanes to form a plurality of substantially equal volume pump chambers when said divider-disk is in neutral position, said disc being operably driven from said rotor by co-operative interaction between said disc slots and rotor 'vanes, a plurality of peripheral compartments disposed radially outwardly of said pump chambers in said divider-disk, duct means interconnecting one of said peripheral compartments to one of said pump chambers on opposite sides of said divider-disk, duct and valve means to control the flow of fluid to said pump compartments.

3. A variable displacement pump comprising a housing having therein a main, generally spherical chamber, a hollow core rotor positioned in said chamber, a plurality of radial vanes mounted on said rotor, a generally cylindrical divider-disk having a plurality of slots there 15 in to receive'said vanes positioned in said chamber, said divider-disk being pivotally mounted in said housing about a diameter thereof to divide said chamber into two areas of complementary variable capacity pump chambers, a plurality of peripheral compartments disposed radially outwardly of said main chamber in said disk, duct means in said disc interconnecting said pump and peripheral chambers, control means external of said housing for varying the pivotal position of said disk, a ducting core member mounted within the core of said rotor, said core having thereon a plurality of oil channels positioned to sequentially interconnect corresponding pump chambers on opposite sides of said divider-disk, and angle indicating assembly means comprising gear and shaft members inter- 5 connecting said divider-disk and said control means to indicate the angular tilt of said divider-disk about the.

divider-disk being pivotally mounted in said housing about a diameter thereofto divide said chamber into two areas of complementary variable capacity inner compartments, a plurality of peripheral compartments disposed radially outwardly of said main chamber in said disk, duct means interconnecting said inner chamber to the corresponding opposite peripheral chambers, a ducting core member mounted within the core of said rotor, said core having thereon a plurality of oil channels positioned to sequentially interconnect corresponding pump chambers on either side of said divider-disk, a plurality of shallow cavities cut in said divider-disk at the edge of said slots, a plu rality of oil sealing inserts disposed in said shallow cavities V in the edge of said divider-disk, a plurality of springs urging said inserts outwardly of said cavity and into contact with the vanes of said rotor, a plurality of ducts and valve assemblies positioned between adjacent shallow cavities in said divider-disk, said ducts and valve assemblies being adapted to select the high-pressure side of said dividerdisk and to admit pressure therefrom into said sealing insert cavities whereby tight oil sealing contact is maintained, control means mounted externally of said housing for varying the displacement of said pump and angle indicating assembly means mounted on said housing comprising gear and shaft members interconnecting said divider-disk and said control means to indicate the angular tilt of said divider-disk about the pivotal axis thereof.

5. A variable displacement pump comprising a housing forming a generally spherical chamber therein; an inputoutput shaft mounted axially therein; a rotor member mechanically coupled to said shaft; six radial vanes mount ed on said rotor and extending radially outwardly therefrom in a plane passing through the axis of said shaft; a divider disk assembly pivotally mounted about a diameter thereof in said housing, said diameter being disposed at right angles to the axis of said input-output shaft; said assembly comprising a disk support member having an annular groove therein and a barrel portion partially forming an inner compartment within said outer housing about said rotor, and a circular disk member having six radial slots cut therein mounted in said annular groove; said disk assembly forming twelve substantially equal volume pump compartments within said inner chamber when said disk is at right angles to said shaft; twelve peripheral compartments formed on either side of said divider disk in the portion thereof mounted in said annular groove; duct members interconnecting one of said peripheral cornpartments on one side of said disk to a pump compartment on the other side of said disk; inner chamber end members mounted about said shaft to complete the inner chamber about said rotor; a ducting core member mounted within said shaft to sequentially communicate with said inher pump chambers as said rotor is rotated about said shaft,

lti a pluralityof vane members disposed between said inner chamber and said outer housing to form together with said disk support assembly at least two compartments within said'housing duct means connected to said two compartments, duct means for introducing oil into one of said compartments and withdrawing oil from the other whereby the angle of tilt of said disk support assembly may be varied.

6. A variable displacement pump comprising in combination an outer housing having formed therein an inner pumping chamber and an outer control chamber; a work input shaft rotatably mounted in said housing and extending into said pumping chamber; a hollow core rotor member operatively connected to and positioned within said pumping chamber; a plurality of vanes mounted on said rotor; a divider disk member having a corresponding plurality of slots therein; a non-rotating support assembly mounted in said outer housing for pivotal motion about an axis disposed at right angles to the axis of said rotor; said disk support assembly having thereon a pair of flanges positioned to form the dividing wall between said inner pumpingchamber and said outer control chamber and an annular groove therein disposed outwardly of said pumping chamber adapted to receive said divider disc; said divider disk and disk support serving to divide said inner and outer chambers into two areas of complementary variable capacity compartments; said disk being operatively driven from said rotor by co-operative interaction between said disc slots and rotor vanes; friction reducing means disposed about the periphery of said divider disk to at least partially compensate for an hydrostatic unbalance therein; fluid ducting means interconnecting said slots for controlling the flow of fluid in said inner compartments; secondary fluid ducting means interconnecting said outer compartments to control the pivotal position of said disk and support.

7. 'A variable displacement fluid pump comprising in combination an outer housing having formed therein by non-rotating end plates fixed to said housing an inner pumping chamber and an outer control chamber; a work input shaftrotatably mounted in said housing and extending into said pumping chamber; a rotor member operatively connected to said shaft and positioned within said pumping chamber; at least three radial vanes mounted on said rotor; a divider disk member having at least three corresponding; slots therein; a disk support assembly mounted in said outer housing for pivotal motion about an axis disposed at right angles to the axis of said rotor; said disk support assembly having thereon a pair of flanges positioned to form the dividing wall between said inner pumping chamber and said outer control chamber and an annular groove therein disposed outwardly of said pumping chamber adapted to receive said divider disk; said divider disk and a disk support serving to divide said innerand outer chambers into two areas of complementary variable capacity compartments; said disk being operatively driven from said rotor by cooperative interaction betweendisc slots and rotor vanes; a plurality of slots formed in said rotor; fluid ducting means interconnecting said slots for controlling the flowof fluid in said inner compartments; secondary fluid ducting means interconnecting said outer compartments to control the pivotal position of said disk and supports.

8. A variable displacement pump comprising in combination an outer housing having formed therein by nonrotating end plates fixed thereto an inner pumping chamber and an outer control chamber; a work input shaft rotatably mounted in said housing and extending into said pumping chamber; a rotor member operatively connected to said shaft and positioned within said pumping chamber; a plurality of vanes mounted on said rotor; the outermost surface of-each of said vanes containing a shaped groove; a divider disk member having a plurality of slots therein; the outermost end of each of said slots having aoeaaoe assembly mounted in said outer housing for pivotal motion about an axis disposed at right angles to the axis of said rotor; said disk support assembly having thereon a pair of flanges positioned to form the dividing wall between said inner pumping chamber and said outer control chamber and an annular groove therein disposed outwardly of said pumping chamber adapted to receive said divider disk; said divider disk and disk support serving to divide said inner and outer chambers into two areas of complementary variable capacity compartments, said disk being operatively driven from said rotor by cooperative interaction between disc slots and rotor vanes; a plurality of slots formed in said rotor; fluid ducting means interconnecting said slots for controlling the flow of fluid in said inner compartments; secondary fluid ducting means interconnecting said outer compartments to control the pivotal position of said disk and support.

9. A variable displacement pump comprising in combination an outer housing having formed therein by nonrotating end plates fixed thereto an inner pumping chamber and an outer control chamber; a work input shaft rotatably mounted in said housing and extending into said pumping chamber; a rotor member operatively connected to said shaft and positioned within said pumping chamber; a plurality of vanes mounted on said rotor; the outermost surface of each of said vanes containing a shaped groove; a divider disk member having a plurality of slots therein; the outermost end of each of said slots having a radially inwardly extending projection adapted to extend into said shaped grooves each of said slots containing a radial recess in each radial slot face; said recess containing a circumferentially slideable insert; a non-rotating disk support assembly mounted in said outer housing for pivotal motion about an axis disposed at right angles to the axis of said rotor; said disk support assembly having thereon a pair of flanges positioned to form the dividing wall between said inner pumping chamber and said outer control chamber adapted to receive said divider disk; said divider disk and disk support serving to divide said inner and outer chambers into two areas of complementary variable capacity compartments; said disk being operatively driven from said rotor by co-operative interaction between disc slots and rotor vanes; a plurality of slots formed in said rotor; fluid ducting means interconnecting said slots for controlling the flow of fluid in said inner compartments; secondary fluid ducting means interconnecting said outer compartments to control the pivotal position of said disk and support.

References Cited in the file of this patent UNITED STATES PATENTS 951,064 Erickson Mar. 1, 1910 1,020,271 Erickson Mar. 12, 1912 2,242,058 Cuny May 13, 1941 2,318,386 Haines May 4, 1943 2,323,926 McGill July 13, 1943 2,371,228 Dodge Mar. 13, 1945 2,431,122 Iakobsen Nov. 18, 1947 2,443,074 Kraft June 8, 1948 2,691,349 Cuny Oct. 12, 1954 2,808,006 Paulsmeier et al Oct. 1, 1957 2,828,695 Marshall Apr. 1, 1958 FOREIGN PATENTS 1,009,488 Germany May 29, 1957 (KL 59c) 1,123,013 France June 4, 1956 

